Turbocharger and Method

ABSTRACT

A turbocharger includes a turbine, a compressor, and a bearing housing forming a bearing bore. A bearing arrangement is disposed between a shaft interconnecting the turbine and compressor wheels, and the bearing housing. The bearing arrangement includes an outer bearing race element that frictionally engages the bearing bore at both ends. At a first end, a bearing retainer that is connected to the bearing housing includes a cylindrical element that extends into the bearing bore and abuts and frictionally engages a first end of the outer bearing race element. The second end of the outer bearing race element abuts and frictionally engages a stop formed at an end of the bearing bore such that no pins or other structures are required for preventing rotation of the outer bearing race element in the bearing bore.

TECHNICAL FIELD

This patent disclosure relates generally to turbochargers and, moreparticularly, to turbochargers used on internal combustion engines.

BACKGROUND

Internal combustion engines are supplied with a mixture of air and fuelfor combustion within the engine that generates mechanical power. Tomaximize the power generated by this combustion process, the engine isoften equipped with a turbocharged air induction system.

A turbocharged air induction system includes a turbocharger having aturbine that uses exhaust from the engine to compress air flowing intothe engine, thereby forcing more air into a combustion chamber of theengine than a naturally aspirated engine could otherwise draw into thecombustion chamber. This increased supply of air allows for increasedfuelling, resulting in an increased engine power output.

In conventional turbochargers, engine oil is provided to lubricate andcool bearings in the bearing housing that rotatably support aturbocharger shaft that transfers power from the turbine to thecompressor. In addition to cooling and lubrication, the oil providesdampening for shaft and bearing vibrations when provided in thin filmsas it passes though control or bearing surfaces. Such dampening, whichis sometimes referred to as squeeze film dampening, can providevibration dampening but is often insufficient to provide sufficientdampening in cartridge style bearings. Typical designs may suspend thebearing arrangement within a bore in a bearing housing and use a pin toprevent rotation of the bearing assembly along with the turbochargershaft relative to the bearing housing.

SUMMARY

The present disclosure is applicable to turbochargers for use ininternal combustion engines. In one embodiment, a turbocharger includesa turbine, a compressor, and a bearing housing forming a bearing bore. Abearing arrangement is disposed between a shaft interconnecting theturbine and compressor wheels, and the bearing housing. The bearingarrangement includes an outer bearing race element that frictionallyengages the bearing bore at both ends. At a first end, a bearingretainer that is connected to the bearing housing includes a cylindricalelement that extends into the bearing bore and abuts and frictionallyengages a first end of the outer bearing race element. The outer bearingrace element is otherwise free to rotate within the bearing bore under arotational motion that is dampened by viscosity of oil films presentalong bearing surfaces between the outer bearing race element and thebearing bore. In order for oil to gravity scavenge multiple drain holesare located in the outer race to ensure proper oil scavenging at allcircumferential orientations of the bearing outer race.

Therefore, in one aspect, the disclosure describes a turbocharger thatincludes a turbine having a turbine wheel, a compressor having acompressor wheel, and a bearing housing disposed and connected betweenthe turbine and the compressor. The bearing housing forms a bearing borehaving a stop surface at one end. A shaft is rotatably disposed withinthe bearing housing and extends into the turbine and the compressor. Theturbine wheel is connected to one end of the shaft and the compressorwheel is connected to an opposite end of the shaft such that the turbinewheel is rotatably disposed in the turbine and the compressor wheel isrotatably disposed in the compressor. A bearing arrangement is disposedbetween the shaft and the bearing housing, and includes an outer bearingrace element disposed in the bearing bore. The outer bearing raceelement has a hollow cylindrical shape that engages the bearing borefrictionally along a first end and a second end, and an inner bearingrace element, which engages the shaft and is rotatably disposed withinthe outer bearing race element. A bearing retainer has a cylindricalportion extending into the bearing bore and abutting the first end ofthe outer bearing race element such that, during operation, the outerbearing race element is free to rotate within the bearing bore at arotational motion that is dampened by a viscosity of oil present atbearing surfaces disposed between the outer bearing race element and thebearing bore.

In another aspect, the disclosure describes a method for rotatably andsealably supporting a shaft within a bearing housing of a turbocharger.The method includes connecting a turbine wheel at one end of the shaft,forming a first roller bearing by engaging a first plurality of rollingelements in a first inner race formed in an inner bearing race elementand in a first outer race formed in an outer bearing race element,forming a second roller bearing by engaging a second plurality orrolling elements in a second inner race formed in the inner bearing raceelement and in a second outer race formed in the outer bearing raceelement, and engaging the outer bearing race element between a bearingbore formed in the bearing housing and the shaft, which extends throughthe bearing bore, such that the inner bearing race element rotates withthe shaft with respect to the outer bearing race element. The methodfurther includes frictionally engaging the bearing bore with the outerbearing race element along a first end and a second end, and engagingthe shaft with an inner bearing race element, which is rotatablydisposed within the outer bearing race element. A bearing retainerhaving a cylindrical portion extending into the bearing bore andabutting the first end of the outer bearing race element is provided.The method also includes allowing the outer bearing race element torotate within the bearing bore, and dampening a rotation of the outerbearing race element within the bearing bore by providing a squeeze filmof oil along bearing surfaces between the outer bearing race element andthe bearing bore.

In yet another aspect, the disclosure describes an internal combustionengine having a plurality of combustion chambers formed in a cylinderblock, an intake manifold disposed to provide air or a mixture of airwith exhaust gas to the combustion chambers, and an exhaust manifolddisposed to receive exhaust gas from the combustion chambers. The enginefurther includes a turbine having a turbine housing surrounding aturbine wheel, the turbine housing being fluidly connected to theexhaust manifold and disposed to receive exhaust gas therefrom to drivethe turbine wheel, a compressor having a compressor housing thatsurrounds a compressor wheel, the compressor housing being fluidlyconnected to the intake manifold and disposed to provide air thereto,and a bearing housing disposed and connected between the turbine and thecompressor. The bearing housing forms a bearing bore therethrough thataccommodates a shaft interconnecting the turbine wheel and thecompressor wheel to transfer power therebetween. The shaft is rotatablymounted within the bearing housing and extends into the turbine and thecompressor such that the turbine wheel is connected to one end of theshaft and the compressor wheel is connected to an opposite end of theshaft. A bearing arrangement is disposed between the shaft and thebearing housing. The bearing arrangement includes first and secondbearings, each of the first and second bearings formed by a respectivefirst and second plurality of roller elements engaged between arespective first and second inner race and a respective first and secondouter race.

In one embodiment, an outer bearing race element is disposed within thebearing bore and forms the respective first and second outer races, andan inner bearing race element forms the respective first and secondinner races. The outer bearing race element is disposed in the bearingbore and has a hollow cylindrical shape that engages the bearing borefrictionally along a first end and a second end. A bearing retainerhaving a cylindrical portion extends into the bearing bore and abuts thefirst end of the outer bearing race element. The outer bearing raceelement, during operation, is free to rotate within the bearing bore ata rotational motion that is dampened by a viscosity of oil present atbearing surfaces disposed between the outer bearing race element and thebearing bore.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a block diagram of an internal combustion engine in accordancewith the disclosure.

FIG. 2 is an outline view from a side perspective of a turbocharger inaccordance with the disclosure.

FIG. 3 is a fragmented view through a center of the turbocharger shownin FIG. 2.

FIG. 4 is an enlarged detail view of the turbocharger bearings shown inFIG. 3.

FIGS. 5 and 6 are enlarged detailed views of seals at both ends of theshaft of the turbocharger shown in FIG. 3.

FIG. 7 is an illustration of the fragmented view of FIG. 3 showing flowpaths of oil through the bearing housing of the turbocharger shown inFIG. 2.

FIG. 8 is an enlarged detail of FIG. 7.

FIG. 9 is a fragmented view of two turbocharger bearings in accordancewith the disclosure.

FIGS. 10 and 11 are graphical representations of roto-dynamics for aturbocharger in accordance with the disclosure.

FIGS. 12-15 are illustrations of a bearing housing assembly process inaccordance with the disclosure.

DETAILED DESCRIPTION

This disclosure relates to an improved turbocharger used in conjunctionwith an internal combustion engine to promote the engine's efficientoperation and also the robust and reliable operation of theturbocharger. A simplified block diagram of an engine 100 is shown inFIG. 1. The engine 100 includes a cylinder case 104 that houses aplurality of combustion cylinders 106. In the illustrated embodiment,six combustion cylinders are shown in an inline or “I” configuration,but any other number of cylinders arranged in a different configuration,such as a “V” configuration, may be used. The plurality of combustioncylinders 106 is fluidly connected via exhaust valves (not shown) tofirst exhaust conduit 108 and the second exhaust conduit 110. Each ofthe first exhaust conduit 108 and the second exhaust conduit 110 isconnected to a turbine 120 of a turbocharger 119. In the illustratedembodiment, the turbine 120 includes a housing 122 having a gas inlet124, which is fluidly connected to the first exhaust conduit 108 and thesecond exhaust conduit 110 and arranged to receive exhaust gastherefrom. Exhaust gas provided to the turbine 120 causes a turbinewheel (not shown here) connected to a shaft 126 to rotate. Exhaust gasexits the housing 122 of the turbine 120 through an outlet 128. Theexhaust gas at the outlet 128 is optionally passed through other exhaustafter-treatment components and systems such as an after-treatment device130 that mechanically and chemically removes combustion byproducts fromthe exhaust gas stream, and/or a muffler 132 that dampens engine noise,before being expelled to the environment through a stack or tail pipe134.

Rotation of the shaft 126 causes a wheel (not shown here) of acompressor 136 to rotate. As shown, the compressor 136 can be an axial,radial or mixed flow compressor configured to receive a flow of fresh,filtered air from an air filter 138 through a compressor inlet 140.Pressurized air at an outlet 142 of the compressor 136 is routed via acharge air conduit 144 to a charge air cooler 146 before being providedto an intake manifold 148 of the engine 100. In the illustratedembodiment, air from the intake manifold 148 is routed to the combustioncylinders 106 where it is mixed with fuel and combusted to produceengine power.

An EGR system 102, which is optional, includes an EGR cooler 150, whichis also optional, that is fluidly connected to an EGR gas supply port152 of the first exhaust conduit 108. A flow of exhaust gas from thefirst exhaust conduit 108 can pass through the EGR cooler 150 where itis cooled before being supplied to an EGR valve 154 via an EGR conduit156. The EGR valve 154 may be electronically controlled and configuredto meter or control the flow rate of the gas passing through the EGRconduit 156. An outlet of the EGR valve 154 is fluidly connected to theintake manifold 148 such that exhaust gas from the EGR conduit 156 maymix with compressed air from the charge air cooler 146 within the intakemanifold 148 of the engine 100.

The pressure of exhaust gas at the first exhaust conduit 108, which iscommonly referred to as back pressure, is higher than ambient pressure,in part, because of the flow restriction presented by the turbine 120.For the same reason, a positive back pressure is present in the secondexhaust conduit 110. The pressure of the air or the air/EGR gas mixturein the intake manifold 148, which is commonly referred to as boostpressure, is also higher than ambient because of the compressionprovided by the compressor 136. In large part, the pressure differencebetween back pressure and boost pressure, coupled with the flowrestriction and flow area of the components of the EGR system 102,determine the maximum flow rate of EGR gas that may be achieved atvarious engine operating conditions.

An outline view of the turbocharger 119 is shown in FIG. 2, and afragmented view is shown in FIG. 3. In reference to these figures, andin the description that follows, structures and features that are thesame or similar to corresponding structures and features alreadydescribed may be, at times, denoted by the same reference numerals aspreviously used for simplicity. As shown, the turbine 120 is connectedto a bearing housing 202. The bearing housing 202 surrounds a portion ofthe shaft 126 and includes bearings 242 and 243 disposed within alubrication cavity 206 formed within the bearing housing 202. Thelubrication cavity 206 includes a lubricant inlet port 203 and alubricant outlet opening 205 that accommodate a flow of lubricationfluid, for example, engine oil, therethrough to lubricate the bearings242 and 243 as the shaft 126 rotates during engine operation.

The shaft 126 is connected to a turbine wheel 212 at one end and to acompressor wheel 213 at another end. The turbine wheel 212 is configuredto rotate within a turbine housing 215 that is connected to the bearinghousing 202. The compressor wheel 213 is disposed to rotate within acompressor housing 217. The turbine wheel 212 includes a plurality ofblades 214 radially arranged around a hub 216. The hub 216 is connectedto an end of the shaft 126. In the illustrated embodiment, the turbinewheel 212 is connected at the end of the shaft 126 by welding, but othermethods, such as by use of a fastener, may be used to connect theturbine wheel to the shaft. The turbine wheel 212 is rotatably disposedbetween an exhaust turbine nozzle 230 defined within the turbine housing215. The exhaust turbine nozzle 230 provides exhaust gas to the turbinewheel 212 in a generally radially inward and axial direction relative tothe shaft 126 and the blades 214 such that the turbine 120 is a mixedflow turbine, meaning, exhaust gas is provided to the turbine wheel inboth radial and axial directions. Exhaust gas passing over the turbinewheel 212 exits the turbine housing 215 via an outlet bore 234 that isformed in the housing. The outlet bore 234 is fluidly connected to theoutlet 128 (FIG. 1). The exhaust turbine nozzle 230 is fluidly connectedto an inlet gas passage 236 having a scrolled shape and formed in theturbine housing 215. The inlet gas passage 236 fluidly interconnects theexhaust turbine nozzle 230 with the gas inlet 124 (also see FIG. 1). Itis noted that a single, inlet gas passage 236 is shown formed in theturbine housing 215 in FIG. 3, but in alternative embodiments separatedpassages may be formed in a single turbine housing.

In the embodiment shown in FIG. 3, the inlet gas passage 236 wrapsaround the area of the turbine wheel 212 and outlet bore 234 and is opento the exhaust turbine nozzle 230 around the entire periphery of theturbine wheel 212. A cross sectional flow area of the inlet gas passage236 decreases along a flow path of gas entering the turbine 120 via thegas inlet 124 and being provided to the turbine wheel 212 through theexhaust turbine nozzle 230.

A radial nozzle ring 238, which also forms a shroud for the turbinewheel 212, is disposed substantially around the entire periphery of theturbine wheel 212. As will be discussed in more detail in the paragraphsthat follow, the radial nozzle ring 238 is disposed in fluidcommunication with the inlet gas passage 236 and defines the exhaustturbine nozzle 230 around the turbine wheel 212. As shown in FIG. 3, theradial nozzle ring forms a plurality of vanes 246, which are fixed andwhich are symmetrically disposed around the radial nozzle ring 238 andoperate to direct exhaust gas form the inlet gas passage 236 towards theturbine wheel 212. The shape and configuration of the plurality of vanes246 can vary. Flow channels 250 having an inclined shape are definedbetween adjacent vanes in the first plurality of vanes 246. A flowmomentum of gas passing through the flow channels 250 is directedgenerally tangentially and radially inward towards an inner diameter ofthe turbine wheel 212 such that wheel rotation may be augmented.Although the vanes 246 further have a generally curved airfoil shape tominimize flow losses of gas passing over and between the vanes, thusproviding respectively uniform inflow conditions to the turbine wheel,they also provide structural support to a shroud portion of the radialnozzle ring 238. The radial nozzle ring 238, which includes the shroudportion, is connected to the turbine via a plurality of fasteners 252,but other methods can be used. The fasteners 252 engage a heat shield254, which is connected to a turbine flange 256 formed on the bearinghousing 202 with an interference fit and stakes 258.

The bearing housing 202 encloses a portion of the shaft 126, which isrotationally mounted in a bearing bore 260 formed in the bearing housingby bearings 242 and 243. Each of the bearings 242 and 243 includes anouter race 261 that engages an inner diameter surface of the bearingbore 260, rollers, and an inner race 262 that has a generally tubularshape and extends around the shaft 126 along its length. Oil from thelubricant inlet port 203 is provided by an external oil pump to thebearings 242 and 243 during operation via passages 264, from where itwashes over the bearings to cool and lubricate them before collecting inthe lubrication cavity 206 and draining out of the bearing housingthrough the lubricant outlet opening 205.

The bearings 242 and 243 are axially retained within the bearing bore260 by a bearing retainer 266 disposed between a compressor mountingplate 268 formed on the bearing housing 202 and the compressor wheel213. The bearing retainer 266 forms a central opening 270 having aninner diameter that is smaller than an inner diameter of the bearingbore 260 such that, when the bearing retainer 266 is connected to thebearing housing 202, the bearings 242 and 243 are retained within thebearing bore 260. The bearing retainer 266 is fastened to the compressormounting plate 268 by fasteners 272, but other fastening or retentionstructures may be used, and has a cylindrical portion that engages anouter race 261 axially. The engagement between the cylindrical portion(adjacent 288 as shown in FIG. 3) can create, at times, a frictionalengagement between the bearing housing and the outer race via theretaining plate that removes the necessity of using structures tootherwise prevent rotation of the outer race such as by use of aradially extending pin or similar structures. To prevent material wearat this interface, a nitride surface treatment may be applied to theannular end face of the bearing retainer 266 that abuts the outer race261, or a similar anti-wear treatment that can increase materialhardness can be used. In the illustrated embodiment, the outer race 261may in fact rotate relative to the bearing housing, at a rotation ratethat is much slower than the shaft, during operation. Such rotationduring operation may be dampened by the viscosity of the oil film thatis present in the bearing surfaces B1, B2, B3 and B4 (see FIG. 9). Acomponent of this frictional engagement is also provided on a turbineside of the outer race, which abuts a stop surface at the end of thebearing bore 260, as shown in FIG. 6 (stop surface adjacent end of B4),when the turbine is not operating. During operation, a gap of a fewthousandths of an inch may form between the outer race and the stopsurface that permits oil to pass therethrough.

The compressor 136 includes a compressor vane ring 274 that forms vanes276 disposed radially around the compressor wheel 213. The vanes 276fluidly connect a compressor inlet bore 278, which contains thecompressor wheel 213, with a compressor scroll passage 280 that isformed in the compressor housing 217 and that terminates to a compressoroutlet opening 282. Bolts 284 and circular plate segments 286 connectthe turbine housing 215 to the turbine flange 256 and the compressorhousing 217 to the compressor mounting plate 268. A nut 288 engaged onthe shaft 126 retains the shaft 126 within the bearings 242 and 243.

An enlarged detailed view of the bearings 242 and 243 is shown in FIG.4. In this illustration, and in the other illustrations that follow,structures that are the same or similar to structures previouslydescribed herein will be denoted by the same reference numeralspreviously used for simplicity. Accordingly, the first bearing 242,which can also be referred to as the compressor-side bearing, is formedby a plurality of roller elements 302 that are confined in rolling orsliding motion between an outer race groove 304, which is formed in theouter race 261, and an inner race groove 306, which is formed close tothe compressor-side end of the inner race 262. Similarly, the secondbearing 243, which can also be referred to as the turbine-side bearing,is formed by a plurality of roller elements 308 that are confined inrolling or sliding motion between a corresponding outer race groove 310and inner race groove 312.

The outer race 261 forms various features that facilitate operation ofthe turbocharger 119 and also promote a desirable flow of lubricationoil through the bearing housing 202. More specifically, the outer race261 has a generally hollow cylindrical shape that forms an outer wall orouter casing 314. The outer casing 314 forms the outer race grooves 304and 310 at its ends, and encloses a cylindrical space 316 that surroundsthe shaft 126 and inner race 262 during operation. Close to either end,the outer casing 314 forms two oil collection grooves or oil feedgalleys 318, each of which is axially aligned with the passages 264formed in the bearing housing 202 such that, during operation, oilflowing through the passages 264 collects and fills each of the two oilcollection grooves or oil feed galleys 318. Lubrication passages 320extend through the outer casing 314 and fluidly connect each respectiveoil feed galley 318 with the cylindrical space 316 in an area close tothe inner race grooves 306 and 312, and also the outer race grooves 304and 310, to lubricate and cool the bearings 242 and 243 duringoperation. The outer casing 314 further forms drainage openings 322 thatfluidly connect the cylindrical space 316 with the lubrication cavity206 to drain out any oil collecting within the outer race 261.

The outer race 261 contacts the bearing bore 260 along four cylindricalbearing surfaces, each of which has a diameter and axial length along ashaft centerline, C/L, that has been designed and selected for optimalbearing and dampening performance during operation. Accordingly,beginning from the compressor side of the outer race 261, a firstbearing surface B1 has an outer diameter D1 (see FIG. 9) and extendsalong an axial length L1. A second bearing surface B2 has a diameter D2(FIG. 9) and an axial length L2. A third bearing surface B3 has adiameter D3 (FIG. 9) and extends along an axial length L3. Finally, afourth bearing surface B4 has a diameter D4 (FIG. 9) and extends alongan axial length L4. The bearing surfaces are also illustrated in FIG. 9.

Each of the four bearing surfaces B1, B2, B3 and B4 permits a thin filmor a squeeze film diameter of oil therein having a thickness equal to adifference between the inner diameter D of the bearing bore 260 and theouter diameters D1, D2, D3 and D4. As shown, the two bearing surfaces B1and B2 that straddle the compressor-side oil feed gallery 318 have thesame squeeze film diameter (SFD) and are considered together in terms ofaxial length (L1+L2). Similarly, the two turbine-side bearing surfacesB3 and B4 have the same SFD and are considered together in terms ofaxial length (L3+L4). As used herein, SFD is used to refer to thosehollow cylindrical areas between each bearing surface and the bearingbore through which oil passes during operation. The thickness of thecylindrical areas or gaps are referred to as SFD clearance, while thelength of each cylindrical area (the “height” of the cylindrical area)along the centerline of the shaft is referred to as SFD length.

For the compressor side bearing surfaces, B1 and B2, a ratio of the SFDclearance over the diameter, which can be expressed as (Dx−D)/D, isequal to about 0.0021, where “x” is 1 or 2 and denotes D1 or D2. For thesame bearing surfaces, the SFD length over the diameter, which can beexpressed as (L1 or L2)/D, is equal to about 0.300. For the turbine sidebearing surfaces B3 and B4, a ratio of the SFD clearance over thediameter, which can be expressed as (Dx−D)/D, is equal to about 0.0031,where “x” is 3 or 4 and denotes D3 or D4. For the same bearing surfaces,the SFD length over the diameter, which can be expressed as (L3 orL4)/D, is equal to about 0.200. In other words, in the illustratedembodiment, the cylindrical areas through which oil flows duringoperation, which can act to dampen shaft vibrations and otherexcitations, are thinner and longer on the compressor side than on theturbine side, where they are thicker and shorter, thus providingdifferent dampening characteristics.

During operation, oil provided through the passages fills and, to acertain extent, pressurizes the oil feed galleys 318. Oil from the oilfeed galleys 318 is pushed or passes into the SFDs of the bearingsurfaces B1, B2, B3 and B4, such that oil flows out from each oil feedgalley 318 towards the compressor on one side, the turbine on anopposite side, and towards the center of the bearing housing on bothsides. To promote oil flow through the inner bearing surfaces B2 and B3,the oil flowing towards the center of the bearing housing 202 iscollected by drainage grooves 324 (also see FIG. 8), which are formed onan external surface of the outer race 261, and which direct the oil intothe lubrication cavity 206.

The outer race 261 surrounds the inner race 262, which in turn surroundsa portion of the shaft 126. The inner race 262 forms two end portions326 having a reduced diameter portion that engages the ends of the shaft126. The shaft 126 includes a slender portion 328 having a reduced outerdiameter 330, which is smaller than an increased outer diameter 332 atthe ends of shaft 126. The slender portion 328 extends over an axiallength 334. The increased outer diameter 332 of the shaft 126 mates atits ends with a reduced inner diameter 336 of the two end portions 326of the inner race 262.

To provide torsional and bending rigidity to the shaft 126, the innerrace 262 is advantageously flared along a middle portion thereof to forman increased inner diameter 338. The increased inner diameter 338overlaps in an axial direction with the slender portion 328 to increasethe bending stiffness of the combined structure of the shaft 126 andinner race 262 without considerably increasing the overall mass of thesystem. In the illustrated embodiment, to facilitate assembly, the innerrace 262 is formed by two components, a compressor-side cup 340 and aturbine-side cup 342. One of the cups, in this case the turbine-side cup342, forms a ledge and a wall that accepts therein the free, annularface of the compressor-side cup 340. Together, the compressor-side cup340 and the turbine-side cup 342 form the inner race 262 that has acentral, flared portion 344 and two transition portions 346 connectingthe flared portion 344 with the two end portions 326. Smooth orchamfered transitions 350, which avoid stress concentration, areprovided between the end portions, the transition portions 346, and theflared portion 344, as shown in the enlarged detail of FIG. 8. In theillustrated embodiment, each chamfered transition 350, which can beconvex or concave, is formed at the same radius, but different radii canbe used.

An enlarged detail view of an interface between the compressor wheel 213and the shaft 126 is shown in FIG. 5. In this figure, a diagnosticpassage 402 formed in the bearing housing 202 can be seen. Thediagnostic passage 402 is plugged with a plug 404, which can be removedduring service provide access, for example, to the interior of thebearing housing for installation of instrumentation and/or access to theinterior of the bearing housing.

As can also be seen in FIG. 5, a ring seal 406 is disposed to provide asliding seal between an internal, working chamber of the compressor andthe oil cavity of the bearing housing. More specifically, the ring seal406 is disposed in an open channel 408 that, together with an annularsurface 410 on the inner side of the back of the compressor wheel 213,forms a U-shape. The open channel 408 is formed at the end of anextension of the inner race 262 that is disposed on a compressor-side ofthe bearing 242. The ring seal 406 slidably and sealably engages aninner bore 412 of the bearing retainer 266 such that a sliding seal isprovided between the inner race 262 and the bearing retainer 266 thatprovides sealing against leakage of oil from the bearing housing 202into the compressor housing 217. In addition, the ring seal 406 providessealing against pressurized gas from entering the interior of thebearing housing. A bearing retainer seal 414 is disposed between anouter portion of the bearing retainer 266 and the compressor mountingplate 268. It is noted that an interior 348 (FIG. 4) of the inner race262 is expected to be generally free of oil as no entry openings for oilare provided except, perhaps, the interface between the compressor-sidecup 340 and the turbine-side cup 342. In the event of turbochargerfailure, in a condition when the shaft 126 may be pulled towards theturbine housing, the retention nut 288 may be pulled towards andsealably engage a seat 424, to keep the piston rings engaged and retainthe turbine wheel and shaft assembly within the bearing housing.

In the illustrated embodiment, a tortuous path is also provided todiscourage oil flow towards the ring seal 406. As shown, the end of theinner race 262 forms a radially outward extending portion 416 thatslopes away from the shaft 126. The outward extending portion forms anouter tip portion 418 that is shaped as a cylindrical wall extendingtowards the compressor. The bearing retainer 266 forms an inwardlyfacing cylindrical wall 420 that is axially aligned with the outer tipportion 418 and disposed radially inward therefrom such that ameandering or tortuous path 422 is formed therebetween leading up to thering seal 406.

An enlarged detail view of an interface between the turbine wheel 212and the bearing housing 202 is shown in FIG. 5. In this figure, adrainage groove 502 is formed towards an end 504 of the shaft 126 tofacilitate drainage of oil passing through the innermost bearing surfaceB4 into the scavenge oil gallery. To seal against leakage of oil, and toprovide sealing against pressurized gas from entering the interior ofthe bearing housing, two ring seals are provided between the shaft 126and an inner bore 506 of the turbine flange 256. More specifically, afirst ring seal 508 is disposed in a channel 510 formed in the shaft126, and a second ring seal 512 is disposed in a channel 514, which isalso formed in the shaft 126.

During operation, oil from within the bearing housing 202 is discouragedfrom leakage into the working chamber of the turbine by the sliding andsealing contact of the first ring seal 508 and the second ring seal 512with the shaft 126 and the inner bore 506 of the turbine flange 256. Itis noted that, in the event of a failure in the turbocharger duringwhich the shaft 126 may displace towards the turbine, at least the firstring seal 508 can axially displace within the inner bore 506 for apredetermined distance while still maintaining contact therewith toprovide a seal even under a failure mode, to avoid leakage of oil intothe turbine housing. The same sliding tolerance is provided in the eventhe shaft 126 displaces towards the compressor, in which case the secondring seal 512 can displace within the inner bore 506 while stillmaintaining its sealing function. The ring seals shown herein areadvantageously made of a hardened material such as M2 Steel having ayield stress of about 3,247 MPa (471,000 ksi) and can withstandtemperature differences between the ring and surrounding components ofabout 450 deg. F. In each instance, the rings have a rectangular crosssection, but other cross sections can be used, and have a C-shape thatcan be installed in a channel formed in a shaft to provide a spring-loadagainst a sealably sliding surface that the ring engages.

A simplified oil flow diagram is shown in FIG. 7, where the structuresshown in FIG. 4 are used for illustration of the flow paths. In oneembodiment, a main oil flow 519 is provided at the lubricant inlet port203. At a point A, the supply pressure and flow of oil splits into thepassages 264 to reach the oil feed galleys 318. Point B is taken todescribe oil pressure in the oil feed galley 318 disposed on thecompressor side (left side of the figure), and point C is taken todescribe oil pressure in the oil feed galley 318 disposed on the turbineside (right side of the figure). Oil from the oil feed galleys 318passes through the bearing surfaces as previously described, and drainsinto the lubrication cavity 206. For purpose of description, point E istaken in bearing B1, and point F is taken in bearing B4. Table 1 belowillustrates oil flow rates in gallons per minute (GPM) at differentoperating pressures (low, medium and high, depending on engine speed),and temperatures (cold and hot oil), which are representative of typicalengine operating conditions:

TABLE 1 Oil Flow Data (GPM) Hot Oil Hot Oil Cold Oil Point Low PressureMedium Pressure High Pressure A 0.9 1.6 0.040 B 0.2 0.3 0.003 C 0.2 0.30.004 D 0.1 0.2 0.001 E 0.8 0.2 0.001As can be seen from the above table, the larger gap at point E accountsfor more flow of oil towards the turbine, which promotes more effectivecooling. In the above table, hot oil can be anywhere within a normal oiltemperature operating range for an engine such as between 190 and 230deg. F., and cold oil can be anywhere in a cold start engine operatingrange such as between −30 and 0 deg. F. Similarly, low pressure can bebetween 20 and 40 PSI, medium pressure can be between 50 and 75 PSI, andhigh pressure can be between 90 and 120 PSI.

As discussed above, oil passing through the bearing surfaces B1 and B2on the compressor side, and bearing surfaces B3 and B4 on the turbineside (see FIG. 9), help dampen vibrations and imbalances duringoperation. Such imbalances are advantageously controlled by selectingdifferent oil film thicknesses on both sides of the shaft, which controlthe shaft dynamics to have natural vibration frequencies beyond theoperating range of the engine. For example, for an engine operating athigher speeds and loads, the natural vibration frequencies or at leasttheir prevalent harmonics are configured to occur above or below theexpected range of engine operation. In the present embodiment, thedifference between D1 and D2 with D3 and D4 in the bearing surfaces B1,B2, B3 and B4 produce the desired characteristics.

FIGS. 10 and 11 show graphical representations of the vibrationcharacteristics of a turbocharger in accordance with the presentdisclosure, which was operated to sweep shaft rotation speeds using bothhot oil, for example, oil at a normal operating temperature, and coldoil. As can be seen from the above table, the amount of oil flowingthrough the bearing areas, and also its viscosity, will change withtemperature thus yielding different dampening characteristics againstvibration. The vibration characteristics can be quantified from manydifferent aspects, including a shaft displacement as a percentage of thedisplacement measured, observed or expected with respect to the bearingdiameter at the bearing areas, averaged over the four bearing areas.

The results of a shaft speed sweep on shaft displacement using hot oilare shown in FIG. 10, where shaft speed 516, as a percentage of maximumspeed, is plotted along the horizontal axis, and percentage displacement518, expressed in (%), of a displacement distance with respect to thebearing diameter, is plotted along the vertical axis. Two curves areshown, the dashed lines representing a compressor response curve 520 andthe solid line representing a turbine response curve 522. The compressorresponse curve 520 represents a collection of points showing thepercentage displacement 518 of each test point and the correspondingshaft speed 516 over a range of shaft speeds taken at the compressorwheel (e.g., compressor wheel 213, FIG. 3). Similarly, the turbineresponse curve 522 represents a collection of points showing thepercentage displacement 518 of each test point and the correspondingshaft speed 516 over a range of shaft speeds taken at the turbine wheel(e.g., turbine wheel 212, FIG. 3). The same curves plotted against thesame parameters, but for cold oil, are shown in FIG. 11

As can be seen from the graphs in FIGS. 10 and 11, when the lubricatingoil is warm, a peak vibration of just over 2% can occur at thecompressor wheel speed below 10% of the maximum speed, as denoted bypoint 524 on the graph, and at about that same shaft speed, a vibrationwith a much lower displacement percentage of about 0.5% can occur at theturbine wheel, as denoted by point 526. As can be seen by the compressorresponse curve 520 in FIG. 10, the percent displacement over a range ofshaft speeds between 10% and about 85% of maximum speed, which accountsfor most of the engine's operating range, remains constant at less than1% for the compressor wheel. The turbine response curve 522 shows evenbetter vibration profiles of a relatively constant peak displacement ofless than 0.5% over a speed range between 10% and 100% of the maximumspeed.

When the lubricating oil is cold, as shown in FIG. 11, a peak vibrationof about 7% can occur at the turbine wheel at around 50%, as denoted bypoint 532 on the graph, and at about that same shaft speed, a vibrationwith a much lower displacement percentage of about 4.4% can occur at thecompressor wheel, as denoted by point 530. At a speed of about 5%,similar peaks as those seen in the hot oil condition (FIG. 10) can beseen, with the compressor wheel having a peak displacement percentage ofabout 3.5%, as denoted by point 534, and the turbine wheel having a peakdisplacement percentage of about 1%, as denoted by point 526. In bothcases, the peak displacement at the 5% speed with cold oil is aboutdouble that of hot oil.

As the shaft speed increases, still using cold oil (FIG. 11), thepercent displacement over a range of shaft speeds between 55% and about115%, which accounts for most of the engine's operating range, remainsconstant at less than 1% for the turbine wheel. The compressor responsecurve 520 shows even better vibration profiles of a relatively constantpeak displacement of about than 0.5% over a range between 55% and 115%.With these vibration profiles, shaft roto-dynamics is acceptable untilthe oil warms up, and then settles to a low peak displacement of lessthan 1% over the expected engine operating range. It is noted that, onthe graphs of FIGS. 10 and 11, idle engine speed may be about 10% of theranges shown in the chart.

When assembling a turbocharger in accordance with the disclosure, andespecially when putting together an assembly of the bearing housing 202,certain process steps may be carried out using a fixture, as shown inFIGS. 12-15. In FIG. 12, an assembly of the turbine wheel 212 welded toan end of the shaft 126 is mounted on a fixture 602 in a verticalposition with the turbine wheel at the bottom. After the first ring seal508 and the second ring seal 512 (FIG. 6) are installed on the shaft,the bearing housing 202, which has the heat shield 254 alreadyinstalled, is inserted around the shaft 126 until the turbine flange 256rests on a second fixture 604, thus setting a proper distance betweenthe turbine flange 256 and the turbine wheel 212, as shown in FIG. 13.

Various components including the outer race 261, inner race 263 andbearings 242 and 243 are inserted into the bearing bore 260 around theshaft 126 and, after various seals are installed, the bearing retainer266 is assembled to close the bearing housing 202 and set a properconcentricity between the shaft 126 and the bearing bore 260, as shownin FIG. 14. The compressor wheel 213 is then installed on the free endof the shaft 126, as shown in FIG. 15. In the illustrated assemblysequence, the subassembly of the turbine wheel 212 onto the end of theshaft 126 may be rotationally balanced before assembly of the turbine isundertaken such that the shaft can determine the concentricity of theremaining components assembled thereafter, including the compressorwheel 213, to maintain a balanced assembly. As an optional step, theentire assembly may be trim balanced after assembly to reduceimbalances, especially those imbalances that may be present whenoperating with cold oil. Trim balancing may be accomplished by removingmaterial from the compressor wheel at the central hub and/or at the tipsof the compressor blades. To determine the amount of material to beremoved and the location for such removal, the entire assembly may beplaced on a rotation balancing machine. It is further noted that theengagement of the radial seal within the inner bore of the bearingretainer, which helps place the shaft concentrically into the bearingbore, also reduces the amount of material that must be removed tobalance the assembly when compared to turbochargers having a differentsealing arrangement than what is shown herein.

INDUSTRIAL APPLICABILITY

It will be appreciated that the foregoing description provides examplesof the disclosed system and technique. However, it is contemplated thatother implementations of the disclosure may differ in detail from theforegoing examples. All references to the disclosure or examples thereofare intended to reference the particular example being discussed at thatpoint and are not intended to imply any limitation as to the scope ofthe disclosure more generally. All language of distinction anddisparagement with respect to certain features is intended to indicate alack of preference for those features, but not to exclude such from thescope of the disclosure entirely unless otherwise indicated.

Recitation of ranges of values herein are merely intended to serve as ashorthand method of referring individually to each separate valuefalling within the range, unless otherwise indicated herein, and eachseparate value is incorporated into the specification as if it wereindividually recited herein. All methods described herein can beperformed in any suitable order unless otherwise indicated herein orotherwise clearly contradicted by context.

We claim:
 1. A turbocharger, comprising: a turbine that includes aturbine wheel; a compressor that includes a compressor wheel; a bearinghousing disposed and connected between the turbine and the compressor,the bearing housing forming a bearing bore having a stop surface definedat one end; a shaft rotatably disposed within the bearing housing andextending into the turbine and the compressor, wherein the turbine wheelis connected to one end of the shaft and wherein the compressor wheel isconnected to an opposite end of the shaft such that the turbine wheel isrotatably disposed in the turbine and the compressor wheel is rotatablydisposed in the compressor; a bearing arrangement disposed between theshaft and the bearing housing, the bearing arrangement including anouter bearing race element disposed in the bearing bore, wherein theouter bearing race element has a hollow cylindrical shape that engagesthe bearing bore frictionally along a first end, and an inner bearingrace element, which engages the shaft and is rotatably disposed withinthe outer bearing race element; a bearing retainer having a cylindricalportion extending into the bearing bore and abutting the first end ofthe outer bearing race element; wherein the outer bearing race element,during operation, is free to rotate within the bearing bore at arotational motion that is dampened by a viscosity of oil present atbearing surfaces disposed between the outer bearing race element and thebearing bore; and wherein multiple drain holes are formed in the outerbearing race element, which operate to provide sufficient oil scavengingat all circumferential orientations relative to the outer bearing raceelement.
 2. The turbocharger of claim 1, wherein, during operation, theouter bearing race element floats along first, second, third and fourthbearing surfaces, each carrying a respective squeeze film diameter (SFD)of oil that operates to dampen radial vibrations of the shaft and theouter bearing race element and also operates to damped the rotation ofthe outer bearing race element.
 3. The turbocharger of claim 1, whereinthe bearing retainer further forms an inner bore adjacent the first end,which inner bore has an inner diameter that is smaller than a diameterof the bearing bore.
 4. The turbocharger of claim 3, wherein innerbearing race element forms an extension that includes a channel intowhich a ring seal is disposed, wherein the ring seal sealably andslidably engages the inner bore of the bearing retainer to form a seal.5. The turbocharger of claim 1, further including a nut fastening theshaft with the inner bearing race element.
 6. The turbocharger of claim5, wherein a force tending to pull the shaft towards the turbine istransferred, through the shaft, the nut and the inner bearing raceelement to the outer bearing race element and to an interface betweenthe second end and the stop surface.
 7. The turbocharger of claim 1,wherein the inner bearing race element forms a flared portion having anincreased inner diameter with respect to end portions thereof thatengage the shaft.
 8. The turbocharger of claim 7, wherein the shaft isconnected to the inner bearing race element at end portions, the endportions having a first diameter, the shaft further forming a slenderportion between the end portions, the slender portion having a seconddiameter that is less than the first diameter.
 9. The turbocharger ofclaim 8, wherein the increased inner diameter of the inner bearing raceelement overlaps in an axial direction with the slender portion of theshaft.
 10. The turbocharger of claim 1, wherein the inner bearing raceelement is formed by two components, a compressor-side cup and aturbine-side cup, and wherein a nut engages the compressor-side cup tothe shaft.
 11. A method for rotatably and sealably supporting a shaftwithin a bearing housing of a turbocharger, comprising: connecting aturbine wheel at one end of the shaft; forming a first roller bearing byengaging a first plurality of rolling elements in a first inner raceformed in an inner bearing race element and in a first outer race formedin an outer bearing race element; forming a second roller bearing byengaging a second plurality or rolling elements in a second inner raceformed in the inner bearing race element and in a second outer raceformed in the outer bearing race element; engaging the outer bearingrace element between a bearing bore formed in the bearing housing andthe shaft, which extends through the bearing bore, such that the innerbearing race element rotates with the shaft with respect to the outerbearing race element; frictionally engaging the bearing bore with theouter bearing race element along a first end and a second end, andengaging the shaft with an inner bearing race element, which isrotatably disposed within the outer bearing race element; providing abearing retainer having a cylindrical portion extending into the bearingbore and abutting the first end of the outer bearing race element;allowing the outer bearing race element to rotate within the bearingbore; and dampening a rotation of the outer bearing race element withinthe bearing bore by providing a squeeze film of oil along bearingsurfaces between the outer bearing race element and the bearing bore.12. The method of claim 11, wherein, during operation, the outer bearingrace element floats along first, second, third and fourth bearingsurfaces, each carrying a respective squeeze film diameter (SFD) of oilthat operates to dampen radial vibrations of the shaft and the outerbearing race element.
 13. The method of claim 11, wherein the bearingretainer further forms an inner bore adjacent the first end, which innerbore has an inner diameter that is smaller than a diameter of thebearing bore.
 14. The method of claim 13, wherein inner bearing raceelement forms an extension that includes a channel into which a ringseal is disposed, wherein the ring seal sealably and slidably engagesthe inner bore of the bearing retainer to form a seal.
 15. The method ofclaim 11, further including a nut fastening the shaft with the innerbearing race element.
 16. The method of claim 5, wherein a force tendingto pull the shaft towards the turbine is transferred, through the shaft,the nut and the inner bearing race element to the outer bearing raceelement and to an interface between the second end and the stop surface,thus increasing a magnitude of the frictional engagement therebetween.17. The method of claim 11, further comprising stiffening an assemblythat includes the inner bearing race element and the shaft by providinga flared portion having an increased inner diameter on the inner bearingrace element with respect to end portions thereof that engage the shaft.18. The method of claim 17, wherein the shaft is connected to the innerbearing race element at end portions, the end portions having a firstdiameter, the shaft further forming a slender portion between the endportions, the slender portion having a second diameter that is less thanthe first diameter.
 19. The method of claim 18, wherein the increasedinner diameter of the inner bearing race element overlaps in an axialdirection with the slender portion of the shaft.
 20. An internalcombustion engine having a plurality of combustion chambers formed in acylinder block, an intake manifold disposed to provide air or a mixtureof air with exhaust gas to the combustion chambers, and an exhaustmanifold disposed to receive exhaust gas from the combustion chambers,the engine further comprising: a turbine that includes a turbine housingsurrounding a turbine wheel, the turbine housing being fluidly connectedto the exhaust manifold and disposed to receive exhaust gas therefrom todrive the turbine wheel; a compressor that includes a compressor housingthat surrounds a compressor wheel, the compressor housing being fluidlyconnected to the intake manifold and disposed to provide air thereto; abearing housing disposed and connected between the turbine and thecompressor, the bearing housing forming a bearing bore therethrough thataccommodates a shaft interconnecting the turbine wheel and thecompressor wheel to transfer power therebetween; wherein the shaft isrotatably mounted within the bearing housing and extends into theturbine and the compressor such that the turbine wheel is connected toone end of the shaft and the compressor wheel is connected to anopposite end of the shaft; a bearing arrangement disposed between theshaft and the bearing housing, the bearing arrangement including firstand second bearings, each of the first and second bearings formed by arespective first and second plurality of roller elements engaged betweena respective first and second inner race and a respective first andsecond outer race; an outer bearing race element disposed within thebearing bore and forming the respective first and second outer races,and an inner bearing race element forming the respective first andsecond inner races; wherein the outer bearing race element is disposedin the bearing bore and has a hollow cylindrical shape that engages thebearing bore frictionally along a first end and a second end; a bearingretainer having a cylindrical portion extending into the bearing boreand abutting the first end of the outer bearing race element; whereinthe outer bearing race element, during operation, is free to rotatewithin the bearing bore at a rotational motion that is dampened by aviscosity of oil present at bearing surfaces disposed between the outerbearing race element and the bearing bore